Refrigerating apparatus using non-azeotropic mixed refrigerant

ABSTRACT

A refrigerating apparatus using a non-azeotropic mixed refrigerant may include a compressor operable in a continuous operation mode and configured to compress the non-azeotropic mixed refrigerant, a condenser configured to condense the refrigerant compressed by the compressor, an expander configured to expand the refrigerant condensed by the condenser, and an evaporator configured to evaporate the refrigerant expanded by the expander. A pressure difference (ΔP) of the non-azeotropic mixed refrigerant has a value included in a range of 340 kPa&lt;ΔP&lt;624.7 kPa. Therefore, reliability of components, such as a piston, in the refrigerating apparatus using the non-azeotropic mixed refrigerant may be further improved.

TECHNICAL FIELD

A refrigerating apparatus using a non-azeotropic mixed refrigerant is disclosed herein.

BACKGROUND ART

A refrigerating apparatus has a cavity an inner space of which is maintained at a low temperature. In the refrigerating apparatus, a refrigeration cycle is provided so as to maintain the cavity at a low temperature. In the refrigeration cycle, a refrigerant circulates through processes of compression, condensation, expansion, and evaporation.

There are various types of refrigerants. A mixed refrigerant is a refrigerant in which two or more types of refrigerants are mixed. Mixed refrigerants include azeotropic mixed refrigerant and non-azeotropic mixed refrigerant.

The azeotropic mixed refrigerant is a refrigerant that changes phase without changing a composition of a gas phase and a liquid phase, similar to a single refrigerant. An evaporation temperature of the azeotropic mixed refrigerant is constant between an inlet and an outlet of the evaporator.

In the non-azeotropic mixed refrigerant, a refrigerant having a low boiling point evaporates first, and a refrigerant having a high boiling point evaporates later. Therefore, the non-azeotropic mixed refrigerant has different gas phase and liquid phase compositions during evaporation, and the evaporation temperature is low at the inlet of the evaporator and high at the outlet of the evaporator.

The non-azeotropic mixed refrigerant has a gliding temperature difference (GTD), which is a characteristic in which the temperature changes at equal pressure during phase change. When the non-azeotropic mixed refrigerant is used, the temperature rises when evaporation occurs at equal pressure, and conversely, the temperature decreases during condensation at equal pressure. In other words, the gliding temperature difference of the refrigerant occurs when the state changes from a saturated liquid to a saturated gas.

A thermal efficiency of a heat exchanger may be improved using this phenomenon. For example, the non-azeotropic mixed refrigerant may form a Lorentz cycle in which a temperature between a refrigerant and a heat source is balanced, and efficiency may be improved by reducing irreversible heat exchange.

As an existing technique for applying the non-azeotropic mixed refrigerant, applicant proposed a capillary structure of a refrigerating apparatus in Korean Patent Registration No. 0119839, which is hereby incorporated by reference.

DISCLOSURE Technical Problem

Although thermal efficiency may be improved using a non-azeotropic mixed refrigerant, an optimum composition and optimum operating conditions of the non-azeotropic mixed refrigerant suitable for a refrigeration cycle applied to a refrigerating apparatus are unknown.

Technical Solution

According to embodiments disclosed herein, a refrigerating apparatus using a non-azeotropic mixed refrigerant may include a compressor operable in a continuous operation mode and configured to compress a non-azeotropic mixed refrigerant, a condenser configured to condense the refrigerant compressed by the compressor, an expander configured to expand the refrigerant condensed by the condenser, and an evaporator configured to evaporate the refrigerant expanded by the expander. A pressure difference (ΔP) of the non-azeotropic mixed refrigerant may have a value included in a range of 340 kPa<βP<624.7 kPa. Therefore, it is possible to reduce friction occurring during operation of the piston in the compressor. This operation mode may obtain a greater advantage when the compressor is operated in the continuous operation mode during operation of the refrigerating apparatus. The continuous operation mode is an operation mode corresponding to an intermittent operation mode and may indicate a state in which the compressor is continuously operated without turning off the compressor even when a current high interior temperature is within a target temperature range.

A condensing pressure (Pd) of the non-azeotropic mixed refrigerant may have a value included in a range of 393.4 kPa<Pd<745.3 kPa. Thus, the condensing pressure realized by the non-azeotropic mixed refrigerant may be suitably used for the compressor.

An evaporation pressure (Ps) of the non-azeotropic mixed refrigerant may have a value included in a range of 53.5 kPa<Ps<120.5 kPa. Thus, the evaporative pressure realized by the non-azeotropic mixed refrigerant may be suitably used for the compressor.

In the continuous operation mode, the compressor may be operated even when an interior temperature is in a satisfaction temperature range. In the continuous operation mode, due to the high pressure difference, oil circulation and a piston lift pressure of a gas bearing may be more reliably obtained.

When the compressor is a linear compressor, it is possible to reliably perform a frictional force reduction operation in operation of the linear compressor by further utilizing action due to the pressure difference of the non-azeotropic mixed refrigerant. The linear compressor may include a shell provided with a suction portion or inlet, a cylinder provided in the shell to define a refrigerant compression space, a frame coupled to an outer side of the cylinder, a piston provided to be able to reciprocate within the cylinder in an axial direction, a discharge valve movably coupled to the cylinder to selectively discharge a refrigerant compressed in the refrigerant compression space, and a passage that extends into a space between the cylinder and the frame and through which at least a part or portion of the refrigerant discharged from the discharge valve may flow. Lubrication of the piston may be performed more smoothly.

The cylinder may include a cylinder body in which a nozzle portion or nozzle may be formed, and a cylinder flange portion or flange that extends outward from the cylinder body. The frame may include a frame body that surrounds the cylinder body, a recessed portion or recess into which the cylinder flange portion is inserted, and a seating portion or seat that faces a seating surface of the cylinder flange portion. Therefore, an internal configuration of the linear compressor may be firmly supported.

The passage may include a first passage formed between an outer circumferential surface of the cylinder flange portion and an inner circumferential surface of the recessed portion. Thus, it is possible to provide a passage through which the compressed high pressure non-azeotropic mixed refrigerant may be bypassed.

The passage may include a second passage formed between the seating surface of the cylinder flange portion and a seating surface of the frame. Thus, the high pressure non-azeotropic mixed refrigerant passing through the first passage may be guided.

The passage may include a third passage that extends into a space between an outer circumferential surface of the cylinder body and an inner circumferential surface of the frame body. The high pressure non-azeotropic mixed refrigerant may be guided to a plurality of locations in a longitudinal direction of the piston and cylinder.

The linear compressor may be provided with an assembly tolerance. In this case, there is no difficulty in passing the gas-phase non-azeotropic mixed refrigerant.

The linear compressor may include a cylinder provided with a cylinder stepped portion on an inner circumferential surface, a piston arranged to be able to reciprocate with the cylinder and provided with a piston stepped portion on an outer circumferential surface, the piston stepped portion forming a low pressure between the piston stepped portion and the cylinder stepped portion when moving in one or a first direction and forming a high pressure between the piston stepped portion and the cylinder stepped portion when moving in the other or a second direction, an oil suction passage formed to allow oil to flow between the cylinder stepped portion and the piston stepped portion, and an oil discharge passage formed to allow oil between the cylinder stepped portion and the piston stepped portion to be discharged outside of the cylinder. In a case of the linear compressor, oil circulation may be more smoothly performed by a high pressure difference of the non-azeotropic mixed refrigerant. Therefore, reliability of the refrigerating apparatus using the non-azeotropic mixed refrigerant may be improved.

The non-azeotropic mixed refrigerant may include a first hydrocarbon and a second hydrocarbon. The first hydrocarbon may be isobutane, and the second hydrocarbon may be propane. As an optimum gliding temperature difference may be obtained, a high-efficiency refrigerating system may be obtained.

The isobutane may be provided in a weight ratio of 76% isobutane 87%. Minimum compression work of the refrigeration cycle, compatibility of a production facility of a refrigerating system, a low purchase cost of refrigerant, a high safety of the refrigerating system, efficiency of the refrigeration cycle, and convenience of handling the refrigerant may be obtained.

According to embodiments disclosed herein, a refrigerating apparatus using a non-azeotropic mixed refrigerant may include a linear compressor configured to compress a non-azeotropic mixed refrigerant, a condenser configured to condense the non-azeotropic mixed refrigerant compressed by the compressor, an expander configured to expand the non-azeotropic mixed refrigerant condensed by the condenser, and an evaporator configured to evaporate the non-azeotropic mixed refrigerant expanded by the expander. A pressure difference (ΔP) of the non-azeotropic mixed refrigerant may be 340 kPa<βP<624.7 kPa. It is possible to reliably perform a friction reducing action between the piston and the cylinder in the linear compressor.

The linear compressor may include a piston configured to reciprocate, and a cylinder configured to guide the piston. The non-azeotropic mixed refrigerant of high pressure compressed by the piston may be guided to an inner surface of the cylinder to cause an outer surface of the piston to be lifted on the inner surface of the cylinder. In this case, operational reliability of the linear compressor may be increased by a lift pressure according to the high pressure difference of the non-azeotropic mixed refrigerant.

The non-azeotropic mixed refrigerant may include at least two hydrocarbons. The at least two hydrocarbons may include at least one first hydrocarbon selected from an upper group having an evaporation temperature of −12° C. or more at 1 bar and at least one second hydrocarbon selected from a middle group having an evaporation temperature of −50° C. or more and less than −12° C. at 1 bar. A gliding temperature difference may be 4° C. or more. It is possible to increase cycle efficiency of the refrigerating system in which the non-azeotropic mixed refrigerant is used and improve operation stability of the refrigerating system.

A weight ratio of the at least one first hydrocarbon may be 50% or more. Therefore, compression work of the compressor may be optimized.

The linear compressor may include a piston configured to reciprocate to compress the non-azeotropic mixed refrigerant, and a cylinder configured to guide the piston. Oil pumped by a pressure difference of the non-azeotropic mixed refrigerant may be present on a contact surface between the piston and the cylinder. Therefore, abnormal supply of oil may be prevented, and lubrication of the piston may be performed more smoothly.

According to embodiments disclosed herein, a refrigerating apparatus may include a compressor configured to compress a non-azeotropic mixed refrigerant, a condenser configured to condense the refrigerant compressed by the compressor, an expander configured to expand the refrigerant condensed by the condenser, and an evaporator configured to evaporate the refrigerant expanded by the expander and provide cold air to an inner space of the refrigerating apparatus. The compressor may be selectively operated in an intermittent operation mode and a continuous operation mode, and in the continuous operation mode, the compressor may be controlled to operate continuously even when a temperature of an inner space of the refrigerating apparatus is within a target temperature range. In the case of the non-azeotropic mixed refrigerant, a large pressure difference may be obtained. Therefore, components operated during operation of the compressor in the continuous operation mode may operate more reliably without external influence due to frictional force.

Advantageous Effects

According to embodiments disclosed herein, it is possible to obtain a refrigerating apparatus capable of obtaining high efficiency when a non-azeotropic mixed refrigerant is used.

DESCRIPTION OF DRAWINGS

FIG. 1 is a schematic temperature graph of a non-azeotropic mixed refrigerant and air in a counterflow evaporator;

FIG. 2 is a graph showing a temperature difference between an inlet and an outlet of an evaporator and a gliding temperature difference of a non-azeotropic mixed refrigerant according to compositions of isobutane and propane;

FIG. 3A is a graph showing a refrigeration cycle when isobutane is used as a refrigerant; and

FIG. 3B is a graph showing a refrigeration cycle when a non-azeotropic mixed refrigerant is used as a refrigerant;

FIG. 4 is a schematic view of a refrigerating apparatus according to an embodiment;

FIG. 5 a cross-sectional view of a linear compressor applied to a refrigerating apparatus, when a piston is retracted, according to an embodiment;

and

FIG. 6 is a cross-sectional view of a linear compressor applied to a refrigerating apparatus, when a piston is advanced, according to an embodiment;

FIG. 7 is a cross-sectional view of an oilless linear compressor according to an embodiment;

FIG. 8 is a cross-sectional view of a suction muffler according to an embodiment;

FIG. 9 is a view showing a state in which a first filter is coupled to a suction muffler, according to an embodiment;

FIG. 10 is a view showing a configuration around a compression chamber according to an embodiment;

FIG. 11 is an exploded perspective view showing a state in which a cylinder and a frame are coupled to each other, according to an embodiment;

FIG. 12 is an exploded perspective view of a cylinder and a frame according to an embodiment;

FIG. 13 is an exploded perspective view of a frame according to an embodiment;

FIG. 14 is a cross-sectional view showing a state in which a cylinder and a frame are coupled to each other according to an embodiment;

FIG. 15 is a view of a cylinder according to an embodiment;

FIG. 16 is an enlarged cross-sectional view of portion “A” in FIG. 14;

FIG. 17 is a cross-sectional view showing a state in which a frame and a cylinder are coupled to each other according to an embodiment;

FIG. 18 is an enlarged view of portion “B” of FIG. 17;

FIG. 19 is a cross-sectional view showing refrigerant flow of a linear compressor according to an embodiment;

FIG. 20 is a view showing flow of refrigerant discharged from a compression chamber in first and second passages according to an embodiment; and

FIG. 21 is a view showing a flow of a refrigerant in a third passage.

BEST MODE

Hereinafter, embodiments will be described with reference to the accompanying drawings. The embodiments are not limited to the embodiments discussed hereinafter, and those skilled in the art who understand the spirit will be able to easily propose other embodiments falling within the scope by adding, modifying, and deleting components. However, this also falls within the spirit.

In the following description, contents are divided into technical elements and described. First, a process of selecting a type of a non-azeotropic mixed refrigerant will be described.

Selection of Type of Non-Azeotropic Mixed Refrigerant

Refrigerants to be mixed, which are suitable for the non-azeotropic mixed refrigerant, are proposed. As the refrigerant to be mixed, a hydrocarbon-based (HC-based) refrigerant may be selected. Hydrocarbon-based refrigerant is an eco-friendly refrigerant having a low ozone depletion potential (ODP) and a low global warming potential (GWP). The criteria for selecting a refrigerant suitable for the non-azeotropic mixed refrigerant among the hydrocarbon-based refrigerants may be summarized as follows.

First, from a viewpoint of compression work, when a difference (pressure difference (ΔP)) between a condensing pressure (Pd or p1) and an evaporation pressure (Ps or p2) is smaller, the compression work of the compressor is further reduced, which is advantageous for efficiency. Therefore, refrigerants having a low condensing pressure and a high evaporation pressure may be selected. However, considering reliability of compressors, an evaporation pressure of 50 kPa or more may be selected.

Second, from a viewpoint of utilization of production facilities, refrigerants may be selected which have been used in the past for compatibility of existing facilities and components. Third, from a viewpoint of purchase costs of refrigerants, refrigerants obtainable at low cost may be selected. Fourth, from a viewpoint of safety, refrigerants that are not harmful to humans when refrigerant leaks may be selected.

Fifth, from a viewpoint of reducing irreversible loss, reduction of a temperature difference between a refrigerant and cold air so as to increase efficiency of a cycle is desirable. Sixth, from a viewpoint of handling, refrigerants that can be conveniently handled at a time of work and may be conveniently injected by handlers may be selected.

The above criteria for selecting refrigerants is variously applied in selecting the non-azeotropic mixed refrigerant.

Classification and Selection of Hydrocarbons

Based on evaporation temperature (Tv), candidate refrigerants suggested by the National Institute of Standards and Technology are classified into three (upper, middle, and lower) groups in descending order of evaporation temperature. A density of refrigerant is higher as evaporation temperature increases.

A combination of candidate refrigerants capable of exhibiting an evaporation temperature of −20° C. to −30° C. suitable for the environment of refrigerating apparatuses may be selected. Hereinafter, classification of the candidate refrigerants will be described.

The candidate refrigerants are classified into three types based on boundary values of evaporation temperature, that is, −12° C. and −50° C. The candidate refrigerants classified into the three types are shown in Table 1. It can be seen that the classification of the evaporation temperature changes greatly based on the boundary values.

TABLE 1 Evaporation Evaporation temperature temperature Triple point Hydrocarbon (1 bar) (20 bar) temperature No. group name ° C. 1 upper isopentane 27.5 154.7 −159.85 2 1,2- 10.3 124.8 −136.25 butadiene 3 n-butane −0.9 114.5 −138.25 4 butene −6.6 105.8 −185.35 5 isobutane −12 100.7 −159.65 6 middle propadiene −34.7 68.2 −136.25 7 propane −42.4 57.3 −187.71 8 propylene −47.9 48.6 −185.26 9 lower ethane −88.8 −7.2 −182.80 10 ethylene −104 −29.1 −169.15

Referring to Table 1, refrigerants that may be mixed as the non-azeotropic mixed refrigerant may be selected and combined in each region. First, which group is selected among the three groups will be described. There may be one case in which refrigerants are selected from the three groups and three refrigerants are mixed, and three cases in which refrigerants are selected from two groups and two refrigerants are mixed.

When at least one refrigerant is selected from each of the three groups and three or more refrigerants are mixed, the temperature rise and drop in the non-azeotropic mixed refrigerant may be excessively great. In this case, design of the refrigerating system may be difficult.

Thus, the non-azeotropic mixed refrigerant may be obtained by selecting at least one refrigerant from each of two groups. At least one refrigerant may be selected from each of the middle group and the lower group, from each of the upper group and the middle group, and from each of the upper group and the lower group. Among them, a composition in which at least one refrigerant selected from each of the upper group and the middle group is mixed may be provided as the non-azeotropic mixed refrigerant.

When at least one refrigerant selected from each of the middle group and the lower group is mixed, the evaporation temperature of the refrigerant is excessively low. Thus, a difference between interior temperature and the evaporation temperature of the refrigerant is excessively great in a general refrigerating apparatus. Therefore, efficiency of the refrigeration cycle deteriorates and power consumption increases.

When at least one refrigerant selected from each of the upper group and the lower group is mixed, a difference in evaporation temperature between the at least two refrigerants is excessively great. Therefore, unless a special high-pressure environment is created, each refrigerant is classified into a liquid refrigerant and a gaseous refrigerant under actual use conditions. For this reason, it is difficult to inject the at least two refrigerants together into a refrigerant pipe.

Selection of Hydrocarbons in Groups of Hydrocarbons

Which refrigerant is selected from the upper group and the middle group will be described hereinafter.

First, the refrigerant selected from the upper group will be described. At least one refrigerant selected from the upper group may be used as the non-azeotropic mixed refrigerant.

As isopentane and butadiene have a relatively high evaporation temperature, the inner temperature of the evaporator of the refrigerating apparatus is limited and freezing efficiency deteriorates. Isobutane and N-butane may be used without changing components of the refrigeration cycle, such as the compressor of the refrigerating apparatus, currently used. Therefore, their use is most expected among the refrigerants included in the upper group.

N-butane has a smaller compression work than isobutane, but has a low evaporation pressure (Ps), which may cause a problem in the reliability of the compressor. For this reason, isobutane may be selected from the upper group. As described above, selection of at least one from the other hydrocarbons included in the upper group is permissible.

The refrigerant selected from the middle group will be described hereinafter. At least one refrigerant selected from the middle group may be used in the non-azeotropic mixed refrigerant.

As propadiene has a smaller pressure difference (ΔP) than that of propane, efficiency is high. However, propadiene is expensive and harmful to respiratory systems and skin when humans inhale due to leakage. Propylene has a greater pressure difference than that of propane, and thus, compression work of the compressor is increased.

For this reason, propane may be selected from the middle group. As described above, selection of at least one from the other hydrocarbons included in the middle group is permissible.

For reference, isobutane may also be referred to as R600a, and propane may also be referred to as R290. Although isobutane and propane may be selected, other hydrocarbons belonging to the same group may be applied in obtaining properties of the non-azeotropic mixed refrigerant, even where there is no specific mention in the following description. For example, if it is possible to obtain a similar gliding temperature difference of the non-azeotropic mixed refrigerant, other compositions than isobutane and propane may be used.

Selection of Ratio of Selected Hydrocarbon Refrigerant, Considering Power Consumption of Compression Work

As the refrigerant to be mixed in the non-azeotropic mixed refrigerant, isobutane is selected from the upper group and propane is selected from the middle group. Ratios of the refrigerants to be mixed in the non-azeotropic mixed refrigerant may be selected as follows.

Power consumption of the compressor, which is a main energy consumption source of the refrigerating system, depends on pressure difference. In other words, as the pressure difference increases, more compression work need to be consumed. As the compression work increases, efficiency of the cycle further deteriorates.

Isobutane has a smaller pressure difference (ΔP) than that of propane. For this reason, the non-azeotropic mixed refrigerant may be provided with a weight ratio of isobutane of 50% or more and a weight ratio of propane of 50% or less.

In the case of a composition in which the non-azeotropic mixed refrigerant includes isobutane and propane mixed at a ratio of 5:5, the condensing pressure is 745.3 kPa, the evaporation pressure is 120.5 kPa, and the pressure difference is 624.7 kPa. In the case of a composition in which the non-azeotropic mixed refrigerant is substantially isobutane with a very small amount of propane, the condensing pressure is 393.4 kPa, the evaporation pressure is 53.5 kPa, and the pressure difference is 340.0 Pa.

The pressure is obtained by measuring an average value when the compressor is turned on under ISO power consumption measurement conditions. All values related to the composition of the non-azeotropic mixed refrigerant are obtained under the same conditions.

Ranges of the condensing pressure, the evaporation pressure, and the pressure difference of the non-azeotropic mixed refrigerant may be known using a mixing ratio of isobutane to propane that can reduce the compression work as described above.

Selection of Ratio of Selected Hydrocarbon Refrigerant, Considering Irreversible Loss of Evaporator

As described above, the non-azeotropic mixed refrigerant has a gliding temperature difference (GTD) upon phase change. Using the gliding temperature difference, evaporators may be sequentially installed in a freezer compartment and a refrigerating compartment to provide an appropriate temperature atmosphere for each partitioned space. According to the gliding temperature difference, a temperature difference between air and refrigerant evaporated in each evaporator may be reduced, thereby reducing irreversibility occurring during heat exchange. Reduction in irreversible loss may reduce the loss of the refrigerating system.

FIG. 1 is a schematic temperature graph of a non-azeotropic mixed refrigerant and air in a counterflow evaporator. In FIG. 1, the horizontal axis represents progress distance, and the air and the non-azeotropic mixed refrigerant move in opposite directions as indicated by arrows. In FIG. 1, the vertical axis represents temperature. Referring to FIG. 1, 1 is a line for air, 2 is a line for the non-azeotropic mixed refrigerant, 3 is a line for temperature rise of the non-azeotropic mixed refrigerant, 4 is a line for temperature drop of the non-azeotropic mixed refrigerant, and 5 is a line for a single refrigerant. Referring to the line 1 for air, for example, the temperature of the air may drop from a range of −20° C. to −18° C. and the air may pass through the evaporator.

Referring to the line 2 for the non-azeotropic mixed refrigerant, the temperature of the non-azeotropic mixed refrigerant may rise from −27° C. and the non-azeotropic mixed refrigerant may pass through the evaporator. The gliding temperature difference of the non-azeotropic mixed refrigerant may change according to the ratio of isobutane to propane. When the gliding temperature difference is increased, the line 2 for the non-azeotropic mixed refrigerant may move toward the line 3 for the temperature rise of the non-azeotropic mixed refrigerant. When the gliding temperature difference is decreased, the line 2 for the non-azeotropic mixed refrigerant may move toward the line 4 for the temperature drop of the non-azeotropic mixed refrigerant. For reference, as there is no phase change in the single refrigerant, there is no temperature change in the line 5 for the single refrigerant.

Irreversible loss when heat exchange occurs cannot be avoided due to the temperature difference between two interfaces where heat exchange occurs. For example, when there is no temperature difference between interfaces of two objects that exchange heat with each other, there is no irreversible loss, but heat exchange does not occur.

However, there are various methods for reducing irreversible loss due to heat exchange. A representative method is to configure a heat exchanger with counterflow. A counterflow heat exchanger may reduce irreversible loss by allowing the temperature difference between moving fluids to be reduced as much as possible.

In the case of an evaporator to which the non-azeotropic mixed refrigerant is applied, the heat exchanger may be configured with counterflow as shown in FIG. 1. As the temperature of the non-azeotropic mixed refrigerant is increased during evaporation due to the gliding temperature difference, the temperature difference between the air and the non-azeotropic mixed refrigerant may be reduced. When the gliding temperature difference of the non-azeotropic mixed refrigerant and the temperature difference of the air are reduced, irreversible loss may be reduced and efficiency of the refrigeration cycle may be increased.

The gliding temperature difference of the non-azeotropic mixed refrigerant may not be increased infinitely due to limitations of the refrigerant. In addition, when the gliding temperature difference of the non-azeotropic mixed refrigerant is changed, the gliding temperature difference of the cold air is changed. Accordingly, a size of the evaporator is changed and total efficiency of the refrigeration cycle is affected. For example, when the gliding temperature difference is increased, the inlet temperature of the refrigerant is decreased or the outlet temperature of the refrigerant is overheated, thus reducing efficiency of the refrigeration cycle.

On the other hand, the gliding temperature difference of the non-azeotropic mixed refrigerant and the temperature difference of the air may converge to zero if a size of the heat exchanger is infinitely large. However, considering mass productivity and cost reduction of the heat exchanger, in the case of a general refrigerating apparatus, the gliding temperature difference of the non-azeotropic mixed refrigerant and the temperature difference of the air are about 3° C. to 4° C.

FIG. 2 is a graph showing a temperature difference between an inlet and an outlet of an evaporator and a gliding temperature difference of a non-azeotropic mixed refrigerant according to compositions of isobutane and propane. The horizontal axis represents a content of isobutane, and the vertical axis represents a temperature difference.

Referring to FIG. 2, when isobutane and propane are each included in 100%, there is no temperature change while isobutane and propane undergo evaporation as a single refrigerant. When isobutane and propane are mixed, there are the gliding temperature difference of the non-azeotropic mixed refrigerant and the temperature difference between the inlet and the outlet of the evaporator. A temperature difference 11 between the inlet and the outlet of the evaporator is smaller than a gliding temperature difference 12 of the non-azeotropic mixed refrigerant. This may be caused by incomplete heat transfer between the refrigerant and air.

When the gliding temperature difference of the non-azeotropic mixed refrigerant is greater than the temperature difference between the inlet and the outlet of the evaporator, characteristics of the non-azeotropic mixed refrigerant may be well utilized. Also, it is advantageous from a viewpoint of reducing irreversibility in heat exchange and increasing efficiency of the refrigeration cycle. Likewise, the gliding temperature difference of the non-azeotropic mixed refrigerant may be greater than the temperature difference of the air passing through the evaporator.

In a general refrigerating apparatus, the temperature difference of the air passing through the inlet and the outlet of the evaporator may reach 4° C. to 10° C. In most cases, the temperature difference of air is close to 4° C. For this reason, the gliding temperature difference of the non-azeotropic mixed refrigerant may be maintained higher than 4° C. Maintaining the gliding temperature difference to be at least 4.1° C. or higher, which is minimally higher than the temperature difference between the inlet and the outlet of the evaporator, may be advantageous. When the gliding temperature difference of the non-azeotropic mixed refrigerant is less than 4.1° C., thermal efficiency of the refrigeration cycle may decrease.

In contrast, when the gliding temperature difference of the non-azeotropic mixed refrigerant is greater than 4.1° C., the temperature difference between the air and the refrigerant at the outlet side of the refrigerant decreases, irreversibility decreases, and thermal efficiency of the refrigeration cycle increases. That the temperature difference between the air and the refrigerant at the outlet side of the refrigerant decreases means that the line 2 for the non-azeotropic mixed refrigerant moves toward the line 3 for the temperature rise of the non-azeotropic mixed refrigerant in FIG. 1.

In FIG. 2, when the gliding temperature difference of the non-azeotropic mixed refrigerant is 4.1° C., isobutane is 90%, and when the gliding temperature difference of the non-azeotropic mixed refrigerant is greater than 4.1° C., isobutane is less than 90%. In order to minimize compression work of the compressor, isobutane may be 50% or more.

As a result, a weight ratio of the non-azeotropic mixed refrigerant provided as isobutane and propane may be expressed as in Equation 1.

50%≤isobutane≤90%  [Math FIG. 1]

Propane is the remaining or other component in the weight ratio of the non-azeotropic mixed refrigerant.

As the gliding temperature difference of the non-azeotropic mixed refrigerant increases, irreversible loss may be reduced. However, when the gliding temperature difference is excessively great, a size of the evaporator becomes excessively large in order to secure a sufficient heat exchange passage between the refrigerant and the air. A space inside of the refrigerating apparatus may be secured when the evaporator applied to a general household refrigerating apparatus is designed with a capacity of 200 W or less. For this reason, the gliding temperature difference of the non-azeotropic mixed refrigerant may be limited to 7.2° C. or less.

In addition, when the gliding temperature difference of the non-azeotropic mixed refrigerant is excessively great, the temperature of the inlet of the evaporator may be too low or the outlet of the evaporator outlet may be overheated too quickly, based on the non-azeotropic mixed refrigerant. An available area of the evaporator may be reduced and efficiency of the heat exchange may decrease.

At the outlet of the evaporator, the temperature of the non-azeotropic mixed refrigerant has to be higher than the temperature of the air introduced into the evaporator. Otherwise, efficiency of the heat exchanger decreases due to reversal of the temperatures of the refrigerant and air. When this condition is not satisfied, efficiency of the refrigerating system may be reduced.

In FIG. 2, when the gliding temperature difference of the non-azeotropic mixed refrigerant is 7.2° C., isobutane is 75%, and when the gliding temperature difference of the non-azeotropic mixed refrigerant is less than 7.2° C., isobutane is more than 75%. As a result, considering this condition and the condition of Equation 1 together, a weight ratio of the non-azeotropic mixed refrigerant provided as isobutane and propane may be expressed as in Equation 2.

75%≤isobutane≤90%  [Math FIG. 2]

Propane is the remaining or other component in the weight ratio of the non-azeotropic mixed refrigerant.

Selection of Ratio of Selected Hydrocarbon Refrigerant, Considering Compatibility of Production Facilities and Components

The temperature difference between the inlet and the outlet of the evaporator of a general refrigerating apparatus may be set to 3° C. to 5° C. This is due to various factors, such components of the refrigerating apparatus, internal volume of the machine room, heat capacity of each component, and size of the fan, for example. When a composition ratio of the non-azeotropic mixed refrigerant capable of providing the temperature of the inlet and the outlet of the evaporator, that is, 3° C. to 5° C., is found in FIG. 2, it can be seen that isobutane is between 76% and 87%.

As a result of the above discussion, the non-azeotropic mixed refrigerant that satisfies all of the above-described conditions may be expressed as Equation 3.

76%≤isobutane'87%  [Math FIG. 3]

Propane is the remaining or other component in the weight ratio of the non-azeotropic mixed refrigerant.

Ratio of hydrocarbon refrigerant to be finally applied

The isobutane application range that can be selected on the basis of the various criteria described above may be determined to be 81% to 82%, which is the middle range of Equation 3. Propane may occupy the remaining portion or component of the non-azeotropic mixed refrigerant.

The case of using only isobutane was compared with the case of using the non-azeotropic mixed refrigerant in which 85% of isobutane and 15% of propane were applied. In both cases, the evaporators were constructed in parallel to form the cycle of the refrigerating system.

The experimental conditions were −29° C. and −15° C. and the inlet temperatures of the compressors were 25° C., respectively. Due to the difference in the refrigerant, the temperature of the condenser was 31° C. when using only isobutane and 29° C. when using the non-azeotropic mixed refrigerant.

FIGS. 3A and 3B are tables for comparison of the refrigeration cycle in each case. FIG. 3A is a graph showing the refrigeration cycle when only isobutane is used. FIG. 3B is a graph showing the refrigeration cycle when the non-azeotropic mixed refrigerant is used.

In the experiment according to FIGS. 3A-3B, it can be seen that when the non-azeotropic mixed refrigerant is used, improvement in coefficient of performance was approximately 4.5%.

FIG. 4 is a schematic view of a refrigerating apparatus according to an embodiment. Referring to FIG. 4, a refrigerating apparatus according to an embodiment may include a machine room 631, a freezer compartment 632, and a refrigerating compartment 633. The refrigerating apparatus forms a refrigeration cycle that operates the non-azeotropic mixed refrigerant. In the refrigeration cycle, a compressor 621 that compresses the refrigerant, an expander 622 that expands the compressed refrigerant, a condenser 623 that condenses the expanded refrigerant, and evaporators 624 and 625 may be included.

The compressor 621, the expander 622, and the condenser 623 may be provided in the machine room 631. The first evaporator 624 may be provided in the freezer compartment 632. The second evaporator 625 may be provided in the refrigerating compartment 633. The freezer compartment and the refrigerating compartment may be referred to as an “interior space”.

The non-azeotropic mixed refrigerant is lower in the first evaporator 624 than in the second evaporator 625. As the first evaporator 624 is placed in the freezer compartment 632, the refrigeration cycle may be operated more appropriately in the partitioned space of the refrigerating apparatus. Therefore, irreversible loss may be further reduced in an evaporation operation of the evaporator 625.

Hereinafter, a linear compressor applicable as compressor 621 provided in the refrigerating apparatus will be described.

FIG. 5 a cross-sectional view of a linear compressor applied to a refrigerating apparatus, when a piston is retracted, according to an embodiment. FIG. 6 is a cross-sectional view of a linear compressor applied to a refrigerating apparatus, when a piston is advanced, according to an embodiment.

In the linear compressor according to this embodiment, as shown in FIGS. 5 and 6, a linear compressor 60 is installed so as to be buffered inside a shell 50. The shell 50 may include a lower shell 51 with an open upper portion and an upper shell 52 mounted to cover an upper side of the lower shell 51. An airtight space may be defined between the lower shell 51 and the upper shell 52, and oil O may be contained in an inner lower portion of the lower shell 51.

The oil allows a piston to be lubricated at an interface of a cylinder. When lubrication is not smooth, it may cause a big problem in reliability of the linear compressor.

The shell 50 may include a suction pipe 53 through which fluid may be suctioned that passes therethrough, and a discharge pipe 54 through which the fluid compressed by linear compressor 60 may be discharged that passes therethrough. The linear compressor 60 may include a damper 55 installed in the lower shell 51 to provide vibrational support. The linear compressor 60 may include a cylinder block 66 provided with a cylinder 62, a back cover 74 provided with a fluid suction pipe 72, a piston 80 provided to linearly reciprocate in the cylinder 62 and including a fluid suction passage 78 and a suction port 79 formed such that the fluid is suctioned into the cylinder 62, a suction valve 82 installed in the piston 80 to open and close the fluid suction passage 78, a linear motor 84 that linearly reciprocates the piston 80, and a discharge valve assembly 92 that forms a compression chamber C between the discharge valve assembly 92 and the piston 80. When the fluid inside the compression chamber C is compressed to a predetermined pressure or higher, the compressed fluid is discharged to the discharge pipe 54, which may be a loop pipe.

The cylinder 62 may be installed at a center of the cylinder block 66. The back cover 74 may be mounted on a stator cover 152, which is described hereinafter, as a coupling member, such as a coupling bolt.

A flange 81 may be formed at a rear end of the piston 80. The flange 81 may be connected to the linear motor 84 by a coupling member, such as a coupling bolt, to receive drive force of the linear motor 84.

The suction valve 82 may be an elastic member coupled to a front end surface of the piston 80 by the coupling bolt and open and close the suction port 79 by a pressure difference between the compression chamber C and the suction port 79.

The linear motor 84 may include an outer core 85 installed in the cylinder block 64, a bobbin 86 installed in the outer core 85, a coil 87 wound around the bobbin 86, an inner core 88 installed in the cylinder block 64 so as to have a predetermined gap with the outer core 85, a magnet 89 located between the outer core 85 and the inner core 88 so as to reciprocate linearly by the electromagnetic force formed by the coil 87, and a magnet frame 90 in which the magnet 89 is mounted and which is coupled to the flange 81 of the piston 80 to transmit a linear motion to the piston 80.

The discharge valve assembly 92 may include a discharge valve 93 that opens and closes a front end of the cylinder 62, an inner discharge cover 96 in which the discharge valve 93 may be elastically supported by discharge spring 94 and a fluid discharge hole 95 may be formed, an outer discharge cover 97 in which a passage may be formed between the outer discharge cover 97 and the inner discharge cover 96, and a connection pipe 98 mounted on the outer discharge cover 97 and connected to the discharge pipe 54.

The linear compressor 60 may be provided with an oil suction passage 10 such that the oil O contained in the shell 50 may flow between the cylinder 62 and the piston 80. An oil discharge passage 20 may be provided such that the oil O between the cylinder 62 and the piston 80 may be discharged. As shown in FIG. 5 when the piston 80 is retracted, a low pressure is formed between the cylinder 62 and the piston 80. As shown in FIG. 6, a pump 30 that forms a high pressure is provided between the cylinder 62 and the piston 80.

As described above, the pump 30 may be operated by a pressure difference between a low pressure and a high pressure formed between the cylinder 62 and the piston 80. When the non-azeotropic mixed refrigerant is used, the pressure difference between the low pressure and the high pressure formed between the cylinder and the piston is large. Therefore, more oil may be pumped with a greater force, as compared to a refrigerating apparatus using a single refrigerant provided with equal cooling power.

The description of the non-azeotropic mixed refrigerant and the pump 30 will be described hereinafter with operation of the linear compressor.

The oil suction passage 10 may include an oil pipe 11 immersed in the oil O contained in the shell 50 and mounted on the cylinder block 66, an oil cover 13 that communicates with the oil pipe 11 and having an oil passage 12 formed between the oil cover 13 and the cylinder block 66, a cylinder block suction passage 14 formed in the cylinder block 66 such that oil suctioned into the oil passage 12 passes through the cylinder block 66, and a cylinder suction passage 15 formed in the cylinder 62 such that oil suctioned into the cylinder block suction passage 14 is suctioned into the inside of the pump 30. The oil discharge passage 20 may include a cylinder discharge passage 21 formed in the cylinder 62 such that the oil O inside of the pump 30 may be discharged, and a cylinder block discharge passage 22 formed in the cylinder block 66 such that the oil O discharged to the cylinder discharge passage 21 may be discharged through the cylinder block 66.

The pump 30 may include a cylinder stepped portion or step 31 formed on an inner circumferential surface of the cylinder 62, and a piston stepped portion or step 32 formed on an outer circumferential surface of the piston 80 to form a low pressure between the piston stepped portion 32 and the cylinder stepped portion 31 when the piston 80 is retracted, and form a high pressure between the piston stepped portion 32 and the cylinder stepped portion 31 when the piston 80 is advanced. An inclined portion of the cylinder stepped portion 31 faces an inclined portion of the piston stepped portion 32. An inner diameter D1 in front of the inclined portion of the cylinder 62 is smaller than an outer diameter D2 behind the inclined portion of the piston 80. The cylinder stepped portion 31 and the piston stepped portion 32 form a cylindrical space when the piston 80 is retracted.

The linear compressor 60 may further include an oil suction valve 40 that opens one or a first side of the oil suction passage 10 by the low pressure formed in the pump 30 when the piston 80 is retracted, and seals the one side of the oil suction passage 10 when the low pressure is released. The oil suction valve 40 may be made of an elastic member fixed to the cylinder block 66 by a coupling member, such as a coupling bolt, and a portion thereof may be bent to open an inlet of the oil suction passage 10, in particular, the oil passage 12.

The linear compressor 60 may further include an oil discharge valve 150 that opens one or a first side of the oil discharge passage 20 by the high pressure formed in the pump 30 when the piston 80 is retracted, and seals the one side of the oil discharge passage 20 when the high pressure is released. The oil discharge valve 150 may be made of an elastic member fixed to the cylinder block 66 by a coupling member, such as a coupling bolt, and a portion thereof may be bent to open an outlet of the oil discharge passage 20, in particular, the cylinder block discharge passage 22.

Reference numeral 152 represents a stator cover coupled to the outer core 86 by a coupling member, such as a coupling bolt, to cover a rear surface of the outer core 85. Reference numeral 27 represents a spring supporter, in which a first spring 29 may be disposed between the spring supporter 27 and the back cover 72, and a second spring 28 may be disposed between the spring supporter 27 and the stator cover 152. The spring supporter 27 may be fixed to the flange 81 of the piston 80 by a coupling member, such as a coupling bolt. Reference numeral 160 represents a muffler installed on a rear end side of the piston 80 to guide the fluid suctioned into the suction pipe 71 of the back cover 72 to the fluid suction passage 78 of the piston 80 and reduce noise.

Operation of the linear compressor configured as described above will be described hereinafter.

First, when a voltage is applied to the coil 87, a magnetic field is formed around the coil 87. The magnet 89 linearly reciprocates due to interaction with the magnetic field. The linear reciprocating motion of the magnet 89 may be transmitted to the piston 80 through the magnet frame 90, and the piston 80 may reciprocate linearly in the cylinder 62.

The suction valve 82 and the discharge valve 93 may be opened or closed by a pressure difference before and after compression according to the linear reciprocating motion of the piston 80. The fluid inside of the shell 50 may be suctioned into the compression chamber C by sequentially passing through the fluid suction pipe 72 of the back cover 74, the muffler 160, the fluid suction passage 78, and the suction port 79 of the piston 80. The fluid may be compressed by the piston 80 and discharged by sequentially passing through the discharge valve assembly 92 and the discharge pipe 54.

As described above, the piston 80 linearly reciprocates, and while the fluid inside of the shell 50 is suctioned/compressed/discharged, the oil O contained in the inner lower portion of the shell 50 may be suctioned into the pump 30 according to a pressure change inside of the pump 30, lubricate/cool between the cylinder 62 and the piston 80, and be discharged outside of the linear compressor 60.

The pressure change of the pump 30 and the oil supply process according to the pressure change will be described hereinafter.

When the piston 80 is retracted as shown in FIG. 5, the piston stepped portion 32 is positioned away from the cylinder stepped portion 31. A low pressure is formed between the piston stepped portion 32 and the cylinder stepped portion 31. The oil suction valve 40 is partially bent by the low pressure to open the inlet of the oil suction passage 10, in particular, the oil passage 12. The oil discharge valve 150 may seal the oil discharge passage 20, in particular, the cylinder block discharge passage 150 by the low pressure. The oil O contained in the inner lower portion of the shell 50 may pass through the oil pipe 11, the oil passage 12, the cylinder block suction passage 14, and the cylinder suction passage 15 due to the low pressure, be suctioned into the space between the piston stepped portion 32 and the cylinder stepped portion 31, and lubricate/cool the cylinder 62 and the piston 80.

On the other hand, when the piston 80 is advanced as shown in FIG. 6, the piston stepped portion 32 may be positioned close to the cylinder stepped portion 31. A high pressure may be formed between the piston stepped portion 32 and the cylinder stepped portion 31. The oil suction valve 40 may seal the inlet of the oil suction passage 10, in particular, the oil passage 12, due to the high pressure. The oil discharge valve 150 may be partially bent to open the oil discharge passage 20, in particular, the cylinder block discharge passage 150. The oil in the space between the piston stepped portion 32 and the cylinder stepped portion 31 may pass through the cylinder discharge passage 21 and the cylinder block discharge passage 22 due to the high pressure and be discharged outside of the linear compressor 60.

The pump 30 may supply a large amount of oil with a greater force as the pressure difference of the linear compressor increases. The pressure difference of the linear compressor may correspond to a pressure difference (ΔP) between an evaporation pressure and a condensation pressure of the refrigerant.

When the non-azeotropic mixed refrigerant is used as the working fluid of the linear compressor, it can operate with a greater pressure difference (ΔP) than when using isobutene alone, which is widely used as a single refrigerant. The pressure difference of the non-azeotropic mixed refrigerant applied to embodiments will be described hereinafter.

First, in the case of a composition in which isobutane and propane are mixed at a ratio of 5:5, the condensing pressure is 745.3 kPa, the evaporation pressure is 20.5 kPa, and the pressure difference is 624.7 kPa. In the case of a composition in which the non-azeotropic mixed refrigerant is substantially isobutane and a very small amount of propane is mixed in, the condensing pressure is 393.4 kPa, the evaporation pressure is 53.5 kPa, and the pressure difference is 340.0 Pa.

As a result, the pressure difference of the non-azeotropic mixed refrigerant of this embodiment may have a range of 340.0 kPa or more and 624.7 kPa or less. As such, when the non-azeotropic mixed refrigerant is used as the working fluid, oil may be supplied to the linear compressor with a greater pressure difference. In this case, it is possible to prevent a phenomenon in which the oil supply passage is blocked, oil circulation is slowed, or a contact surface between the piston and the cylinder is overheated.

When the non-azeotropic mixed refrigerant is used in the linear compressor, a greater advantage may be expected when the refrigerating system is operated in a continuous operation mode. First, the continuous operation mode will be described hereinafter.

The refrigerating apparatus is driven differently in the refrigeration cycle according to a temperature region divided by an upper limit reference value and a lower limit reference value based on a set or predetermined temperature of an interior of the refrigerating apparatus. More specifically, the temperature region is divided into three types as follows. The temperature region may be divided into a first case in which the interior temperature is within a satisfaction temperature range falling between the upper and lower limit reference values, a second case in which the interior temperature is in a dissatisfaction temperature range exceeding the upper limit reference value, and a third case in which the interior temperature is in a supercooling temperature range in which the interior temperature is below the lower limit reference value.

A controller of the refrigerating apparatus may control the refrigerating apparatus to operate to supply cooling power when the interior temperature reaches the dissatisfaction temperature range, and control the refrigerating apparatus to stop supplying cooling power when the interior temperature reaches the supercooling temperature range. Such an operation mode may be referred to as an “intermittent operation mode”.

When the upper limit reference value and the lower limit reference value are set narrowly such that an amount of temperature change in the refrigerating apparatus is reduced, a freshness of an article may be improved and a storage period of the article may be improved. Further, a constant temperature function may be improved and the refrigerating apparatus may operate more suitably.

In the intermittent operation mode, the number of times the refrigeration cycle is driven and stopped increases to improve the constant temperature function. Due to this, reliability of components is reduced by frequent on/off of refrigeration cycle components, and power consumption increases each time the refrigerating apparatus is switched from off to on. Also, when it is switched from off to on, excessive cooling power is initially supplied, and the constant temperature function deteriorates.

In order to solve this problem, the continuous operation mode may be applied to reduce the number of times the refrigeration cycle is driven and stopped. In the intermittent operation mode, when the interior temperature reaches the upper limit reference value and the lower limit reference value, the controller of the refrigerating apparatus performs a switching control such that the refrigerating apparatus starts or stops supplying cooling power. When the interior temperature is in the satisfaction temperature range, the controller controls the refrigerating apparatus to stop supplying cooling power.

On the other hand, in the continuous operation mode, the cooling power supply of the refrigerating apparatus is not stopped even when the interior temperature is within the satisfaction temperature range. The controller may control the refrigerating apparatus based on a current temperature measured by a temperature sensor and a target temperature. For example, the controller may control the cooling power of the refrigerating apparatus based on the difference between the target temperature and the current temperature detected by the temperature sensor. As another example, the controller may control the cooling power of the refrigerating apparatus based on the increase or decrease in the interior temperature detected by the temperature sensor at regular time intervals.

The controller may control the cooling power of the refrigerating apparatus based on the cooling power of the refrigerating apparatus previously determined. For example, the controller may determine the cooling power to be output at a current time point as a value changing proportional to a sum of a first output and a second output of the refrigerating apparatus previously determined. As such, as the refrigerating apparatus is operated in the continuous operation mode, the interior temperature of the refrigerating apparatus may be maintained near the target temperature without deviating to the dissatisfaction temperature region.

The continuous operation mode may be performed in a state in which an article is in a frozen or refrigerated state, and a special operation may be performed in the intermittent operation mode according to a special situation. For example, when a start condition of the special operation in the refrigerating apparatus is satisfied, the controller of the refrigerating apparatus may be switched to the intermittent operation mode to drive the refrigerating apparatus, and when an end condition of the special operation is satisfied, the controller may switch to the continuous operation mode to control operation of the refrigerating apparatus. The special operation may include a defrosting operation, a door load response operation, or an initial power application operation, for example.

The defrosting operation will be described as an example. When a defrosting cycle inside of the refrigerating apparatus has elapsed, and/or when the evaporator temperature in the refrigerating apparatus reaches a preset or predetermined specific value, the controller may end the continuous operation mode and switch to the intermittent operation mode. Operation of the refrigerating apparatus in the defrosting operation will be described hereinafter.

First, in order to pre-cool the internal temperature of the refrigerating apparatus that will rise during the defrosting operation, the controller may perform a deep cooling operation of controlling the refrigerating apparatus to supply a greater cooling power than that in the previous continuous operation mode. When a predetermined time has elapsed after the deep cooling operation or the temperature reaches a set or predetermined temperature, the controller may end the deep cooling operation.

When the deep cooling operation ends, the controller may stop supplying the cooling power of the refrigerating apparatus, and the defrosting heater may be turned on to melt ice formed on the evaporator. In this manner, the defrosting operation of supplying heat to the refrigerating apparatus may be performed.

When the defrosting operation is completed, the interior temperature of the refrigerating apparatus increased during the defrosting operation has to be rapidly lowered. The controller may allow the evaporator to receive a greater cooling power than that received in the previous continuous operation mode. This may be referred to as a “post-defrosting operation”.

A point at which the post-defrosting operation ends is the end condition of the defrosting operation. When the defrosting operation ends or after the defrosting operation ends, the controller may control the refrigerating apparatus to end the intermittent operation mode and switch to the continuous operation mode.

The door load response will be described as another example.

When a predetermined time has elapsed after a door of the refrigerating apparatus is opened or closed and/or when the interior temperature of the refrigerating apparatus has reached a preset or predetermined specific value after the door of the refrigerating apparatus is opened or closed, a start condition of the door load response operation is satisfied. At this time, the controller controls the refrigerating apparatus to end the continuous operation mode and switch to the intermittent operation mode.

Operation of the refrigerating apparatus in the door load response operation will be described hereinafter.

First, in order to remove a heat load introduced into the interior of the refrigerating apparatus by the opening of the door, the controller may perform the door load response operation of controlling the refrigerating apparatus to supply a greater cooling power than that in the previous continuous operation mode. When a predetermined time has elapsed after the door load response operation, or when the interior temperature of the refrigerating apparatus reaches the preset or predetermined temperature, the controller ends the door load response operation. When the door load response operation ends or after the door load response operation ends, the controller may control the refrigerating apparatus to end the intermittent operation mode and switch to the continuous operation mode.

The initial power application operation will be described as another example.

When power is applied to the refrigerating apparatus again after the power of the refrigerating apparatus is cut off, a start condition of the initial power application operation is satisfied. At this time, the controller may control the refrigerating apparatus to operate in the intermittent operation mode rather than the continuous operation mode.

More specifically, in order to rapidly lower the interior temperature of the refrigerating apparatus increased while the power is cut off, the controller performs an initial power application operation of controlling the evaporator to supply a greater cooling power than that in the previous continuous operation mode. When a predetermined time has elapsed after the initial power application operation, or when the interior temperature of the refrigerating apparatus reaches the preset or predetermined temperature, the controller ends the initial power application operation.

When the initial power application operation ends or after the initial power application operation ends, the controller may control the refrigerating apparatus to end the intermittent operation mode and switch to the continuous operation mode. When operating in the continuous operation mode, the controller of the refrigerating apparatus may control the evaporator to receive lower cooling power than that received in the intermittent operation mode. In particular, in the continuous operation mode that is performed for a longer time than the intermittent operation mode, the lower cooling power has to be supplied while being changed.

Such a refrigerating system that variably controls low cooling power may be advantageous when the evaporation temperature of the refrigerant is low. The relationship between the refrigerating system and the evaporation temperature of the refrigerant will be described.

Controlling the controller of the refrigerating apparatus such that the cooling power of the evaporator is lowered means reducing output of the compressor to reduce a flow rate or flow amount of refrigerant circulating inside of a refrigerant pipe. In a situation in which the refrigerant circulates continuously with a low cooling power in order to perform the continuous operation mode, in the refrigerating system in which the evaporation temperature of the refrigerant is relatively high, a circulation speed of the refrigerant in the refrigerating system in which the evaporation temperature of the refrigerant is relatively low has to be high in flow rate or flow amount of the refrigerant. This is because when a same amount of refrigerant evaporates at a lower temperature, more cooling power may be generated.

Consequently, compressor output in the refrigerating system having a high evaporation temperature has to be greater than compressor output in the refrigerating system having a low evaporation temperature. Similarly, there is a disadvantage in that power consumption for driving the compressor is increased. In addition, as a period in which the continuous operation mode is performed is longer, wear of components, such as the compressor, may occur.

Due to the above advantages, it is advantageous that the refrigerating apparatus to which the continuous operation mode is applied uses the non-azeotropic mixed refrigerant mixed, in which a refrigerant, such as propane, having a relatively low evaporation temperature is mixed, rather than isobutane, which has a relatively high evaporation temperature. In the refrigerating apparatus to which the continuous operation mode is applied, reliability problems may occur when the compressor is configured to cause the piston of the compressor to be lifted in the cylinder by oil. As seen in linear compression, linear compressors lubricated with oil cause oil inside of the compressor to circulate due to the pressure difference (ΔP) of the refrigerant generated during a reciprocating motion of the piston. Therefore, as at least one of a frequency or stroke of the linear compressor is lowered during low cooling power driving, the circulating speed of oil and the circulating amount of oil decreases, and thus, an oil supply passage may be blocked.

In the continuous operation mode, as there are many more periods in which low cooling power is supplied than periods in which high cooling power is supplied, problems may occur in the circulation of the oil. In order to solve this problem, when the non-azeotropic mixed refrigerant having a large pressure difference (ΔP) is used, it is possible to reduce the decrease in the circulation speed of the oil and reduce the blocking of the oil supply passage, as compared with a single refrigerant, such as isobutane.

It has been described that the non-azeotropic mixed refrigerant may be applied to the circulating operation of oil for lubrication of the linear compressor. In addition, the non-azeotropic mixed refrigerant may be applied even in the case of a linear compressor in which oil is not used for lubrication. In this case, the linear compressor may be lubricated by air. More precisely, contact between the piston and the cylinder may be prevented by air. Such a compressor may be referred to as an “oilless linear compressor”.

The oilless linear compressor may form a gas layer between the cylinder and the piston instead of oil required to prevent wear and damage between the cylinder and the piston. The oilless compressor may lift the piston in the cylinder due to the pressure difference (ΔP) of the refrigerant generated during reciprocation of the piston. Therefore, it is advantageous to apply the non-azeotropic mixed refrigerant having a large pressure difference (ΔP).

FIG. 7 is a cross-sectional view of an oilless linear compressor according to an embodiment. Referring to FIG. 7, the linear compressor 100 according to this embodiment may include a substantially cylindrical shell 101, a first cover 102 coupled to one or a first side of the shell 101, and a second cover 103 coupled to the other or a second side of the shell 101. For example, the linear compressor 100 may be laid in a horizontal direction, the first cover 102 may be coupled to a first or right side of the shell 101, and the second cover 103 may be coupled to a second or left side of the shell 101. The linear compressor 100 exemplifies a linear compressor in which oil is not used. The first cover 102 and the second cover 103 may be understood as components of the shell 101.

The linear compressor 100 may include a cylinder 120 provided in the shell 101, a piston 130 that linearly reciprocates in the cylinder 120, and a motor assembly 140 as a linear motor that applies a drive force to the piston 130. When the motor assembly 140 is driven, the piston 130 may reciprocate at a high speed. An operating frequency of the linear compressor 100 according to this embodiment may be, for example, approximately 100 Hz.

The linear compressor 100 may include a suction inlet 104, through which refrigerant may be introduced, and a discharge outlet 105, through which the refrigerant compressed in the cylinder 120 may be discharged. The suction outlet 104 may be coupled to the first cover 102, and the discharge outlet 105 may be coupled to the second cover 103.

The refrigerant suctioned through the suction outlet 104 may flow through a suction muffler 150 to the piston 130. In a process in which the refrigerant passes through the suction muffler 150, noise may be reduced. The suction muffler 150 may be configured by combining a first muffler 151 and a second muffler 153. At least a portion of the suction muffler 150 may be located inside of the piston 130.

The piston 130 may include a piston body 131 having an approximately cylindrical shape, and a piston flange 132 that extends from the piston body 131 in a radial direction. The piston body 131 may reciprocate inside of the cylinder 120, and the piston flange 132 may reciprocate outside of the cylinder 120.

The piston 130 may be made of a non-magnetic aluminum material, for example, aluminum or an aluminum alloy. As the piston 130 is made of the aluminum material, a magnetic flux generated in the motor assembly 140 may be transmitted to the piston 130 while preventing leakage from the piston 130 to the outside. The piston 130 may be formed by a forging method, for example.

The cylinder 120 may be made of a non-magnetic aluminum material, for example, aluminum or an aluminum alloy. A material composition ratio of the cylinder 120 and the piston 130, that is, a type and component ratio may be the same.

As the cylinder 120 is made of the aluminum material, the magnetic flux generated in the motor assembly 200 may be transmitted to the cylinder 120 while preventing leakage from the cylinder 120 to the outside. The cylinder 120 may be formed by an extrusion rod processing method, for example.

As the piston 130 and the cylinder 120 may be made of the same material (aluminum), a coefficient of thermal expansion may be the same. During operation of the linear compressor 100, an environment of high temperature (about 100° C.) may be created inside of the shell 100. As the coefficient of thermal expansion of the piston 130 and the cylinder 120 may be the same, the piston 130 and the cylinder 120 may be thermally deformed by a same amount. As a result, the piston 130 and the cylinder 120 may be thermally deformed in different sizes or directions, thereby preventing interference with the cylinder 120 during operation of the piston 130.

The cylinder 120 may be configured to accommodate at least a portion of the suction muffler 150 and at least a portion of the piston 130. Inside the cylinder 120, a compression space P in which the refrigerant is compressed by the piston 130 is formed. A suction hole 133 through which the refrigerant may be introduced into the compression space P may be defined on or at a front side of the piston 130, and a suction valve 135 that selectively opens the suction hole 133 may be disposed on or at a front side of the suction hole 133. A coupling hole to which a predetermined coupling member may be coupled may be defined in or at an approximately central portion of the suction valve 135.

A discharge cover 160 that defines a discharge space 160 a or a discharge passage for the refrigerant discharged from the compression space P, and a discharge valve assembly 161, 162, and 163 coupled to the discharge cover 160 to selectively discharge the refrigerant compressed in the compression space P may be provided at a front side of the compression space P. The discharge valve assembly 161, 162, and 163 may include discharge valve 161 which may be opened when the pressure of the compression space P is above a discharge pressure to introduce the refrigerant into the discharge space, valve spring 162 provided between the discharge valve 161 and the discharge cover 160 to apply elastic force in an axial direction, and stopper 163 that limits a deformation amount of the valve spring 162. The compression space P may be understood as a space defined between the suction valve 135 and the discharge valve 161.

The “axial direction” may be understood as a direction in which the piston 130 reciprocates, that is, the horizontal direction in FIG. 7. Also, in the “axial direction”, a direction from the suction inlet 104 toward the discharge outlet 105, that is, a direction in which the refrigerant flows may be defined as a “frontward direction”, and a direction opposite to the frontward direction may be defined as a “rearward direction”. On the other hand, the “radial direction” may be understood as a direction perpendicular to the direction in which the piston 130 reciprocates, that is, the vertical direction in FIG. 3.

The stopper 163 may be seated on the discharge cover 160, and the valve spring 162 may be seated on or at a rear side of the stopper 163. The discharge valve 161 may be coupled to the valve spring 162, and a rear portion or a rear surface of the discharge valve 161 may be supported on a front surface of the cylinder 120. The valve spring 162 may include, for example, a plate spring. The suction valve 135 may be disposed on or at one or a first side of the compression space P, and the discharge valve 161 may be disposed on the other or a second side of the compression space P, that is, an opposite side of the suction valve 135.

While the piston 130 linearly reciprocates within the cylinder 120, when the pressure of the compression space P is below the discharge pressure and a suction pressure, the suction valve 135 may be opened to suction the refrigerant into the compression space P. On the other hand, when the pressure of the compression space P is above the suction pressure, the suction valve 135 may compress the refrigerant of the compression space P in a state in which the suction valve 135 is closed.

When the pressure of the compression space P is above the discharge pressure, the valve spring 162 may be deformed to open the discharge valve 161. The refrigerant may be discharged from the compression space P into the discharge space 160 a of the discharge cover 160.

The refrigerant flowing through the discharge space of the discharge cover 160 may be introduced into a loop pipe 165. The loop pipe 165 may be coupled to the discharge cover 160 and extend to the discharge outlet 105. The loop pipe 165 may guide the compressed refrigerant of the discharge space to the discharge outlet 105. For example, the loop pipe 165 may have a shape wound in a predetermined direction, extending roundly, and coupled to the discharge outlet 105.

The linear compressor 100 may further include a frame 110. The frame 110 may fix the cylinder 120 and may be coupled to the cylinder 120 by a separate coupling member. The frame 110 may surround the cylinder 120. That is, the cylinder 120 may be accommodated in the frame 110. The discharge cover 172 may be coupled to a front surface of the frame 110.

At least a portion of the gas refrigerant among the high pressure gas refrigerant discharged through the open discharge valve 161 may flow toward an outer circumferential surface of the cylinder 120 and a space between the cylinder 120 and the frame 110 which are coupled to each other. The refrigerant may be introduced into the cylinder 120 through a gas inlet (122 in FIG. 16) and a nozzle (123 in FIG. 16) formed in the cylinder 120. The introduced refrigerant may flow into the space between the piston 130 and the cylinder 120 such that an outer circumferential surface of the piston 130 is spaced apart from an inner circumferential surface of the cylinder 120. Therefore, the introduced refrigerant may function as a “gas bearing” that reduces friction between the cylinder 120 and the piston 130 during reciprocating motion.

The motor assembly 140 may include outer stators 141, 143, and 145 fixed to the frame 110 and disposed to surround the cylinder 120, an inner stator 148 spaced inward from the outer stators 141, 143, and 145, and a permanent magnet 146 disposed in a space between the outer stator 141, 143, and 145 and the inner stator 148. The permanent magnet 146 may linearly reciprocate by mutual electromagnetic force between the outer stators 141, 143, and 145 and the inner stator 148. The permanent magnet 146 may be provided as a single magnet having one polarity or may be provided by coupling a plurality of magnets having three different polarities.

The permanent magnet 146 may be coupled to the piston 130 by a connecting member 138. The connecting member 138 may be coupled to the piston flange 132 and bent and extended toward the permanent magnet 146. When the permanent magnet 146 reciprocates, the piston 130 may reciprocate together with the permanent magnet 146 in the axial direction.

The motor assembly 140 may further include a fixing member 147 that fixes the permanent magnet 146 to the connecting member 138. The fixing member 147 may be configured, for example, by mixing a glass fiber or a carbon fiber with a resin. The fixing member 147 may surround inner and outer sides of the permanent magnet 146, such that a coupled state between the permanent magnet 146 and the connecting member 138 may be firmly maintained.

The outer stators 141, 143, and 145 may include coil winding bodies 143 and 145, and stator core 141. The coil winding bodies 143 and 145 may include bobbin 143, and coil 145 wound in a circumferential direction of the bobbin 143. A cross-section of the coil 145 may have a polygonal shape, for example, a hexagonal shape. The stator core 141 may be formed, for example, by laminating a plurality of laminations in a circumferential direction, and may surround the coil winding bodies 143 and 145.

A stator cover 149 may be disposed on one or a first side of each of the outer stators 141, 143, and 145. The outer stators 141, 143, and 145 may each have one or a second side supported by the frame 110 and the other side supported by the stator cover 149.

The inner stator 148 may be fixed to an outer circumference of the frame 110. In the inner stator 148, a plurality of laminations may be laminated in the circumferential direction outside the cylinder 120.

The linear compressor 100 may further include a supporter 137 that supports the piston 130, and a back cover 170 spring-coupled to the supporter 137. The supporter 137 may be coupled to the piston flange 132 and the connecting member 138 by a predetermined coupling member.

At a front of the back cover 170, a suction guide 155 may be provided. The suction guide 155 may guide refrigerant suctioned through the suction inlet 104 into the suction muffler 150.

The linear compressor 100 may further include a plurality of springs 176 adjusted in natural frequency to allow the piston 130 to perform a resonant motion. The plurality of springs 176 may include a first spring supported between the support 137 and the stator cover 149, and a second spring supported between the support 137 and the back cover 170.

The linear compressor 100 may further include plate springs 172 and 174 provided on or at both sides of the shell 101 such that internal components of the compressor 100 may be supported by the shell 101. The plate springs 172 and 174 include first plate spring 172 coupled to the first cover 102 and second plate spring 174 coupled to the second cover 103. For example, the first plate spring 172 may be fitted to a portion at which the shell 101 and the first cover 102 are coupled, and the second plate spring 174 may be disposed to fit in a portion at which the shell 101 and the second cover 103 are coupled.

FIG. 8 is a cross-sectional view of a suction muffler according to an embodiment. FIG. 9 is a view showing a state in which a first filter is coupled to the suction muffler, according to an embodiment.

Referring to FIGS. 8 and 9, the suction muffler 150 may include first muffler 151, second muffler 153 coupled to the first muffler 151, and a first filter 310 supported by the first muffler 151 and the second muffler 153. In the first muffler 151 and the second muffler 153, a flow space through which the refrigerant may flow is defined. The first muffler 151 may extend from an inside of the suction inlet 104 toward the discharge outlet 105, and at least a portion of the first muffler 151 may extend into the suction guide 155. The second muffler 153 may extend from the first muffler 151 into the piston body 131.

The first filter 310 may be installed in the flow space to filter foreign matter. The first filter 310 may be made of a magnetic material, such that foreign matter included in the refrigerant, in particular, metal dirt, may be easily filtered. For example, the first filter 310 may be made of stainless steel and may have a predetermined magnetic property, and a rust phenomenon may occur. As another example, the first filter 310 may be coated with a magnetic material or may be configured such that a magnetic is attached to a surface of the first filter 310.

The first filter 310 may be a mesh type filter having a plurality of filter holes and may have a substantially disc shape. The filter holes may each have a diameter or a width of a predetermined size or less. For example, the predetermined size may be about 25 μm.

The first muffler 151 and the second muffler 153 may be assembled by a press-fitting method, for example. The first filter 310 may be assembled by being fitted into press-in portions of the first muffler 151 and the second muffler 153.

A groove 151 a to which at least a portion of the second muffler 153 may be coupled may be formed in the first muffler 151. The second muffler 153 may include a protrusion 153 a inserted into the groove 151 a of the first muffler 151.

The first filter 310 may be supported by the first and second mufflers 151 and 153 in a state in which both sides of the first filter 310 are disposed between the groove 151 a and the protrusion 153 a. In a state in which the first filter 310 is located between the first and second mufflers 151 and 153, when the first muffler 151 and the second muffler 153 move in a direction closer to each other and are pressed, side or outer edge portions of the first filter 310 may be fixed by being disposed between the groove 151 a and the protrusion 153 a.

As such, by providing the first filter 310, foreign matter having a predetermined size or more in the refrigerant suctioned through the suction inlet 104 may be filtered by the first filter 310. Therefore, foreign matter is not included in the refrigerant acting as the gas bearing between the piston 130 and the cylinder 120, thereby preventing the refrigerant from flowing into the cylinder 120. In addition, as the first filter 310 is firmly fixed to the press-in portions of the first and second mufflers 151 and 153, separation from the suction muffler 150 may be prevented.

In this embodiment, it has been described that the groove 151 a is formed in the first muffler 151 and the protrusion 153 a is formed in the second muffler 153. However, the protrusion may be formed in the first muffler 151 and the groove may be formed in the second muffler 153.

FIG. 10 is a view showing a configuration around a compression chamber according to an embodiment. FIG. 11 is an exploded perspective view showing a state in which the cylinder and the frame are coupled to each other, according to an embodiment. FIG. 12 is an exploded perspective view showing a configuration of the cylinder and the frame, according to an embodiment. FIG. 13 is an exploded perspective view of the frame according to an embodiment. FIG. 14 is a cross-sectional view showing a state in which the cylinder and the frame are coupled to each other, according to an embodiment.

Referring to FIGS. 10 to 14, in the linear compressor 100, at least a portion of the refrigerant compressed and discharged in the compression chamber P flows into a space between the frame 110 and the cylinder 120. The space between the frame 110 and the cylinder 120 is understood as a gap formed between an inner surface of the frame 110 and the outer surface of the cylinder 120 by an assembly tolerance of the frame 110 and the cylinder 120. The space between the frame 110 and the cylinder 120 may include passages 410, 420, and 430. The passages 410, 420 and 430 may include first passage 410, second passage 420, and third passage 430, which may be sequentially formed in a direction in which the refrigerant flows.

The cylinder 120 may include an approximately cylindrical cylinder body 121, and a cylinder flange 125 that extends radially from the cylinder body 121. The cylinder body 121 may include gas inlet 122 through which the discharged gas refrigerant may flow. The gas inlet 122 may be formed in a circular shape along an outer circumferential surface of the cylinder body 121.

A plurality of gas inlets 122 may be provided. The plurality of gas inlets 122 may include gas inlets 122 a and 122 b (in FIG. 15) located on one or a first side from an axial center of the cylinder body 121, and gas inlet 122 c (in FIG. 15) located on the other or a second side from the axial center.

The cylinder flange 125 may be provided with a coupling portion 126 coupled to the frame 110. The coupling portion 126 may protrude from an outer circumferential surface of the cylinder flange 125 in an outward direction. The coupling portion 126 may be coupled to cylinder coupling hole 118 of the frame 110 by a predetermined fastening member, for example, a bolt.

The cylinder flange 125 may include a seating surface 127 seated on the frame 110. The seating surface 127 may be a rear portion of the cylinder flange 125 extending radially from the cylinder body 121.

The frame 110 may include a frame body 111 that surrounds the cylinder body 121, and a cover coupling portion 115 that extends in a radial direction of the frame body 111 and coupled to the discharge cover 160. The cover coupling portion 115 may be provided with a plurality of cover coupling holes 116, into which the coupling member coupled to the discharge cover 160 may be inserted, and a plurality of cylinder coupling holes 118, into which the coupling member coupled to the cylinder flange 125 may be inserted The cylinder coupling hole 118 may be defined at a somewhat recessed position from the cover coupling portion 115.

The frame 110 may be provided with a recessed portion or recess 117. The recessed portion 117 may be recessed rearward from the cover coupling portion 115, and the cylinder flange 125 may be inserted into the recessed portion 117. That is, the recessed portion 117 may surround an outer circumferential surface of the cylinder flange 125. A recessed depth of the recessed portion 117 may correspond to a frontward-and-rearward width of the cylinder flange 125.

A predetermined refrigerant flow space, that is, first passage 410 may be formed between an inner circumferential surface of the recessed portion 117 and the outer circumferential surface of the cylinder flange 125. In a state in which the cylinder 120 is assembled to the frame 110, a predetermined assembly tolerance may be formed between the outer circumferential surface of the cylinder flange portion 125 and the inner circumferential surface of the recessed portion 117, and a space corresponding to the assembly tolerance may define the first passage 410.

The high pressure gas refrigerant discharged from the discharge valve 161 may flow through the first passage 410 into second passage 420 provided with a second filter 320. The second filter 320 may be a filter provided between the frame 110 and the cylinder 120 to filter the high pressure gas refrigerant discharged through the discharge valve 161.

A seating portion or seat 113 provided stepwise may be formed at a rear end portion or end of the recessed portion 117. The seating portion 113 may extend radially inward from the recessed portion 117 and be located to face the seating surface 127 of the cylinder flange 125. A ring-shaped second filter 320 may be seated on the seating portion 113.

In a state in which the second filter 320 is seated on the seating portion 113, when the cylinder 120 is coupled to the frame 110, the cylinder flange 125 may press the second filter 320 in front of the second filter 320. That is, the second filter 320 may be disposed and fixed between the seating portion 113 of the frame 110 and the seating surface 127 of the cylinder flange 125.

The second passage 420 may be a passage through which the refrigerant passing through the first passage 410 may flow. A predetermined assembly tolerance may be formed between the seating portion 113 and the seating surface 127 of the cylinder flange portion 125, and the space corresponding to the assembly tolerance may define the second passage 420.

The second filter 320 may be installed in the second passage 420 to prevent foreign matter in the high pressure gas refrigerant flowing through the second passage 420 from flowing into the gas inlet 122 of the cylinder 120, and adsorbs oil contained in the refrigerant.

For example, the second filter 320 may include a nonwoven fabric or an absorbent fabric made of polyethylene terephthalate (PET) fibers. PET has excellent heat resistance and mechanical strength. Further, PET can block foreign matter of 2 μm or more in the refrigerant.

Another embodiment is proposed hereinafter.

In the previous embodiment, the second filter 320 has been described as being installed in the second passage 420. However, the second filter 320 may be installed in the first passage 410, that is, a space between the outer circumferential surface of the cylinder flange 125 and the inner circumferential surface of the recessed portion 117 of the frame 110.

The passages 410, 420 and 430 may include third passage 430 through which the refrigerant passing through the second passage 420 may flow. The third passage 430 may extend rearward from the second passage 420 along the outer circumferential surface of the cylinder body 121 and extend to a space between the rear portion of the frame body 111 and the first body end (121 a in FIG. 15) of the cylinder body 121. The refrigerant flowing through the third passage 430 may flow to the inner circumferential side of the cylinder 120 via the gas inlet 122 and the nozzle 123.

FIG. 15 is a view showing of a cylinder according to an embodiment. FIG. 16 is an enlarged cross-sectional view of portion “A” in FIG. 14.

Referring to FIGS. 15 and 16, the cylinder 120 according to this embodiment may include cylinder body 121 having a substantially cylindrical shape and forming first body end 121 a and second body end 121 b, and cylinder flange 125 that extends radially outward from the second body end 121 b of the cylinder body 121. The first body end 121 a and the second body end 121 b form ends of the cylinder body 121 based on axial center 121 c of the cylinder body 121. The cylinder body 121 may include a plurality of gas inlets 122 through which at least a portion of the refrigerant of the high pressure gas refrigerant discharged through the discharge valve 161 may flow, and a third filter 330 may be installed therein. The cylinder body 121 may further include nozzle 123 that extends radially inward from the plurality of gas inlets 122.

The plurality of gas inlets 122 and the nozzle 123 may be one configuration of third passage 430. Therefore, at least a portion of the refrigerant flowing through the third passage 430 may flow toward the inner circumferential surface of the cylinder 120 through the plurality of gas inlets 122 and the nozzle 123. The plurality of gas inlets 122 may be recessed by a predetermined depth and width from the outer circumferential surface of the cylinder body 121.

The introduced refrigerant may flow between the outer circumferential surface of the piston 130 and the inner circumferential surface of the cylinder 120 and function as a gas bearing. That is, due to the pressure of the refrigerant, the outer circumferential surface of the piston 130 may maintain a state of being spaced from the inner circumferential surface of the cylinder 120.

The plurality of gas inlets 122 may include first gas inlet 122 a and second gas inlet 122 b located at one or a first side from the axial center 121 c of the cylinder body 121, and a third gas inlet 122 c located at the other or a second side from the axial center 121 c. The first and second gas inlets 122 a and 122 b may be located closer to the second body end 121 b based on the axial center 121 c of the cylinder body 121, and the third gas inlet 122 c may be located closer to the first body end 121 a based on the axial center 121 c of the cylinder body 121. That is, the plurality of gas inlets 122 may be disposed in an asymmetrical number based on the axial center 121 c of the cylinder body 121.

Referring to FIG. 15, the internal pressure of the cylinder 120 may be higher at the second body end 121 b closer to a discharge side of the compressed refrigerant than the first body end 121 a closer to a suction side of the refrigerant. Therefore, more gas inlets 122 may be formed toward the second body end 121 b to enhance the function of the gas bearing. However, relatively fewer gas inlets 122 may be formed toward the first body end 121 a.

The cylinder body 121 may further include nozzle 123 that extends from the plurality of gas inlets 122 toward the inner circumferential surface of the cylinder body 121. The nozzle 123 may have a smaller width or size than the gas inlet 122.

A plurality of nozzles 123 may be formed along the gas inlet 122 extending in a circular shape. The plurality of nozzles 123 may be spaced apart from each other.

The nozzle 123 may include an inlet 123 a connected to the gas inlet 122 and an outlet 123 b connected to the inner circumferential surface of the cylinder body 121. The nozzle 123 may be may have a predetermined length from the inlet 123 a toward the outlet 123 b.

A recessed depth and width of the plurality of gas inlets 122, and a length of the nozzle 123 may be determined to be an appropriate size in consideration of a rigidity of the cylinder 120, an amount of the third filter 330, or a magnitude of a pressure drop of the refrigerant passing through the nozzle 123. For example, when the recessed depth and width of the plurality of gas inlets 122 are too large or the length of the nozzle portion 123 is too small, the rigidity of the cylinder 120 may be weakened. When the recessed depth and width of the plurality of gas inlet portions 122 are too small, the amount of the third filter 330 installed in the gas inlet 122 may be too small. When the length of the nozzle 123 is too large, the pressure drop of the refrigerant passing through the nozzle 123 becomes too large. Thus, the refrigerant cannot perform a sufficient function as a gas bearing.

A diameter of the inlet 123 a of the nozzle portion 123 may be larger than a diameter of the outlet 123 b. When the diameter of the nozzle 123 is too large, the amount of the refrigerant flowing into the nozzle 123 among the high pressure gas refrigerants discharged through the discharge valve 161 becomes too large. Thus, a flow rate loss of the compressor is increased. When the diameter of the nozzle 123 is too small, the pressure drop in the nozzle 123 becomes large and performance of the gas bearing decreases. Therefore, in this embodiment, the diameter of the inlet 123 a of the nozzle 123 is relatively large to reduce the pressure drop of the refrigerant flowing into the nozzle 123, and the diameter of the outlet 123 b is formed relatively small such that the amount of the gas bearing flowing through the nozzle 123 may be adjusted to a predetermined value or less.

The third filter 330 may be installed in the plurality of gas inlets 122. The third filter 330 may filter the refrigerant flowing toward the inner circumferential surface of the cylinder 120.

The third filter 330 may prevent foreign matter having a predetermined size or more from being introduced into the cylinder 120 and perform a function for adsorbing oil contained in the refrigerant. The predetermined size may be about 1 μm.

The third filter 330 may include a thread that is wound around the gas inlet 123 a. The thread may be made of a polyethylene terephthalate (PET) material and have a predetermined thickness or diameter.

A thickness or diameter of the thread may be determined to have appropriate dimension in consideration of a strength of the thread. When the thickness or diameter of the thread is too small, the thread may be easily broken due to a very weak strength thereof. On the other hand, if the thickness or diameter of the thread is too large, a filtering effect with respect to foreign matter may be deteriorated due to a very large pore in the gas inlet 122 when the thread is wound. For example, the thickness or diameter of the thread may be formed in hundreds of μm units, and the thread may be configured by combining several tens of μm units of spun threads into multiple strands.

The thread may be wound a number of times and be configured such that an end thereof is fixed with a knot. The number of times the thread is wound may be appropriately selected in consideration of a degree of pressure drop of the gas refrigerant and a filtering effect of foreign matter. When the number of windings is too large, the pressure drop of the gas refrigerant becomes too large, and when the number of windings is too small, foreign matter may not be filtered well.

A tension force of the thread may be formed in an appropriate size in consideration of a deformation degree of the cylinder 120 and a fixing force of the thread. When the tension force is too large, deformation of the cylinder 120 may be caused, and when the tension force is too small, the thread may not be well fixed to the gas inlet 122.

FIG. 17 is a cross-sectional view showing a state in which the frame and the cylinder are coupled to each other according to an embodiment. FIG. 18 is an enlarged view of portion “B” of FIG. 17.

Referring to FIGS. 17 and 18, the linear compressor 100 according to this embodiment may include a sealing pocket 370 that communicates with the third passage 430 and in which a sealing member 350 may be installed. The sealing pocket 370 may be a space in which the sealing member 350 may be installed, and may be formed between an inner circumferential surface of the frame body 111 and the outer circumferential surface of the cylinder body 121. The sealing pocket 370 may be formed at a rear portion of the frame 110 and the cylinder 120. Based on the flow direction of the refrigerant, a flow cross-sectional area of the sealing pocket 370 may be formed to be larger than a flow cross-sectional area of the third passage 430.

The rear portion of the frame body 111 may include a pocket forming portion 112 recessed radially outward from the inner circumferential surface of the frame body 111. The pocket forming portion 112 may form at least one surface of the sealing pocket 370.

The frame main body 111 may further include a second inclined portion 113 that extends obliquely from the pocket forming portion 112 in a rearward inward direction. The cylinder body 121 may include a first inclined portion 128 that forms the sealing pocket 370. The first inclined portion 128 may constitute at least one surface of the sealing pocket 370.

The first inclined portion 128 may extend in an inclined shape rearward from the first body end 121 a of the cylinder body 121. The first inclined portion 128 may extend from an inside of the pocket forming portion 112 to a point corresponding to an inside of the second inclined portion 113.

Due to the recessed structure of the pocket forming portion 112 and the inclined structure of the first inclined portion 128, a radial height of the sealing pocket 370 may be formed to be larger than a diameter of the sealing member 350. An axial length of the sealing pocket 370 may be larger than the diameter of the sealing member 350. That is, the sealing pocket 370 may have a size such that the sealing member 350 is movable without interfering with the frame body 111 or the cylinder body 121.

An interval or distance between a rear portion of the first inclined portion 128 and a rear portion of the second inclined portion 113 may be smaller than the diameter of the sealing member 350. Therefore, when refrigerant flows backward along the third passage 430 between operations of the linear compressor 100, the sealing member 350 may move rearward due to the pressure of the refrigerant and seal the spaced space. As such, as the sealing member 350 may be disposed between the cylinder 120 and the frame 110 to seal the third passage 430, it is possible to prevent the refrigerant in the third passage 430 from leaking outside of the frame 110.

When the sealing member 350 is provided to be movable in the sealing pocket 370 and the compressor is driven such that the refrigerant in the third passage 430 flows, the sealing member 350 may be pressed against the cylinder 120 and the frame 110. Thus, deformation of the cylinder 120 due to a pressing force of the sealing member 350 may be prevented.

Hereinafter, flow of the refrigerant between operations of the linear compressor will be described.

FIG. 19 is a cross-sectional view showing refrigerant flow of a linear compressor according to an embodiment. FIG. 20 is a view showing flow of refrigerant discharged from a compression chamber in first and second passages according to an embodiment. FIG. 21 is a view showing the flow of refrigerant in a third passage.

First, the refrigerant flow in the linear compressor will be described with reference to FIG. 19. Referring to FIG. 19, the refrigerant flows into the inside of the shell 101 through the suction inlet 104 and suction guide 155 to the suction muffler 150. The refrigerant flows into the second muffler 153 through the first muffler 151 of the suction muffler 150 and flows into the piston 130. In this process, suction noise of the refrigerant may be reduced.

As the refrigerant passes through the first filter 310 provided in the suction muffler 150, foreign matter having a predetermined size (25 μm) or more may be filtered. The refrigerant that passes through the suction muffler 150 and exists inside of the piston 130 is suctioned into the compression space P through the suction hole 133 when the suction valve 135 is opened.

When the refrigerant pressure in the compression space P exceeds the discharge pressure, the discharge valve 161 is opened and the refrigerant is discharged to the discharge space of the discharge cover 160 through the open discharge valve 161. The discharge valve 161 is moved forward and spaced apart from the front surface of the cylinder 120. In this process, the valve spring 162 is elastically deformed forward. The stopper 163 limits the amount of deformation of the valve spring 162 to a certain extent. The refrigerant discharged into the discharge space 160 a of the discharge cover 160 flows to the discharge outlet 105 through the loop pipe 165 coupled to the discharge cover 160 and is discharged outside of the compressor 100.

At least a portion of the refrigerant in the discharge space 160 a of the discharge cover 160 may flow through the space between the cylinder 120 and the frame 110, that is, the first passage 410 and the second passage 420. The refrigerant may be filtered by the second filter 320 in the process of flowing through the first passage 410 or the second passage 420.

The filtered refrigerant may flow toward the outer circumferential surface of the cylinder body 121 through the third passage 430, and at least a portion of the refrigerant may flow into the plurality of gas inlets 122 formed in the cylinder body 121. The refrigerant introduced into the gas inlet 122 may be filtered by the third filter 330 and introduced into the cylinder 120 through the nozzle 123.

The refrigerant introduced into the cylinder 120 may flow between the inner circumferential surface of the cylinder 120 and the outer circumferential surface of the piston 130 and act to separate the piston 130 from the inner circumferential surface of the cylinder 120 (gas bearing). As such, the high pressure gas refrigerant may be bypassed into the cylinder 120 to provide lift pressure to the reciprocating piston 130 and may act as a bearing at the interface between the cylinder and the piston. Therefore, wear between the piston 130 and the cylinder 120 may be reduced. As an oil bearing is not used, friction loss due to oil may not be generated even when the compressor 100 is operated at high speed.

The refrigerant flowing to the inner surface of the cylinder and the outer surface of the piston may be a compressed refrigerant and have a condensing pressure (Pd). In contrast, the space in which the refrigerant flows has an evaporating pressure (Ps) before being compressed. Therefore, as the difference between the condensing pressure and the evaporation pressure is greater, the piston 130 may be lifted from the inner surface of the cylinder 120 with a greater force.

In addition, as a plurality of filters is provided in the passage of the refrigerant flowing inside of the compressor 100, it is possible to remove foreign matter contained in the refrigerant. This may improve reliability of the refrigerant to act as the gas bearing. Therefore, it is possible to prevent a phenomenon in which abrasion is generated in the piston 130 or the cylinder 120 by foreign matter contained in the refrigerant.

As oil contained in the refrigerant is removed by the plurality of filters, it is possible to prevent friction loss caused by oil. The first filter 310, the second filter 320, and the third filter 330 may be collectively referred to as a “refrigerant filter device” in that the first filter 310, the second filter 320, and the third filter 330 filter the refrigerant to act as the gas bearing.

The refrigerant flowing through the third passage 430 acts on the sealing member 350. That is, the pressure of the refrigerant acts on the sealing member 350, and the sealing member 350 is moved from the sealing pocket 370 to a point between the first inclined portion 128 of the cylinder 120 and the second inclined portion 113 of the frame 110.

The sealing member 350 is in close contact with the cylinder 120 and the frame 110 and seals the space between the cylinder 120 and the frame 110, for example, the space between the first inclined portion 128 and the second inclined portion 113. Therefore, it is possible to prevent refrigerant in the third passage 430 from leaking to the outside through the space between the cylinder 120 and the frame 110.

When driving of the linear compressor 100 is stopped, the pressure of the refrigerant acting on the sealing member 350 is released. Thus, the contact force between the sealing member 350, the cylinder 120, and the frame 110 is weakened. As a result, the sealing member 350 is freely movable within the sealing pocket 220, for example, in a state of being spaced apart from the first inclined portion 128 and the second inclined portion 113 (indicated by a dashed line).

According to this action, the sealing member 350 is in close contact with the cylinder 120 and the frame 110 only when the compressor 100 is driven, and thus, sealing of the third passage 430 may be performed. Therefore, the force applied to the cylinder 120 from the sealing member 350 may be reduced. Therefore, deformation of the cylinder 120 may be prevented.

As the sealing member 350 may be placed in a movable state in the sealing pocket 370, it is possible to prevent interference of the sealing member 350 when the cylinder 120 and the frame 110 are assembled. As a result, assembling of the cylinder 120 and the frame 110 may be facilitated.

As described above, as the non-azeotropic mixed refrigerant is used, the lift pressure of the piston may be increased, and operation of the gas bearing may be performed smoothly. As a minimum lift pressure for lift of the piston may be provided even in a low cooling power operation period, reliability of the linear compressor may be improved. For example, as the lift pressure is increased and friction at the contact surface between the cylinder and the piston decreases, it can advantageously act to increase efficiency of the linear compressor through mechanical design, such as the gas discharge hole and the filter. It is possible to maximize the effect by applying the non-azeotropic mixed refrigerant to the oilless linear compressor in the refrigerating apparatus that performs the continuous operation mode in which the low cooling power operation is frequently performed.

In comparison with a linear compressor using oil, when performing long time low cooling power driving for the continuous operation mode, the oilless linear compressor has an advantage in that the oil is heated by friction between the cylinder and the piston to reduce evaporation and combustion. In the above description, the linear compressor is employed as the compressor. However, embodiments are not limited thereto and may be applied to all compressors in which oil circulation and air bearing are performed due to the pressure difference of the refrigerant during compression of the gas refrigerant.

INDUSTRIAL APPLICABILITY

According to embodiments disclosed herein, as a non-azeotropic mixed refrigerant is used, it is possible to improve durability by preventing wear of components due to a low evaporation temperature of a refrigerant in a continuous operation mode. In addition, it is possible to smoothly operate oil circulation and air bearings for lubrication essential to operation of a compressor. 

1. A refrigerating apparatus including a non-azeotropic mixed refrigerant, the refrigerating apparatus comprising: a compressor operable in a continuous operation mode and to compress the non-azeotropic mixed refrigerant; a condenser to condense the non-azeotropic mixed refrigerant compressed by the compressor; an expander to expand the non-azeotropic mixed refrigerant condensed by the condenser; and an evaporator to evaporate the non-azeotropic mixed refrigerant expanded by the expander, wherein a pressure difference (ΔP) of the non-azeotropic mixed refrigerant is in a range of 340 kPa<ΔP<624.7 kPa.
 2. The refrigerating apparatus according to claim 1, wherein a condensing pressure (Pd) of the non-azeotropic mixed refrigerant is in a range of 393.4 kPa<Pd<745.3 kPa.
 3. The refrigerating apparatus according to claim 1, wherein an evaporation pressure (Ps) of the non-azeotropic mixed refrigerant is in a range of 53.5 kPa<Ps<120.5 kPa.
 4. The refrigerating apparatus according to claim 1, wherein, in the continuous operation mode, the compressor is controlled to operate even when an interior temperature of the refrigerating apparatus is in a target temperature range.
 5. The refrigerating apparatus according to claim 1, wherein the compressor comprises a linear compressor.
 6. The refrigerating apparatus according to claim 5, wherein the linear compressor comprises: a shell including a suction inlet; a cylinder disposed in the shell to define a refrigerant compression space; a frame coupled to an outer side of the cylinder; a piston to reciprocate within the cylinder; a discharge valve movably coupled to the cylinder to selectively discharge the non-azeotropic mixed refrigerant compressed in the refrigerant compression space; and a passage that extends into a space between the cylinder and the frame and through which at least a portion of the non-azeotropic mixed refrigerant discharged from the discharge valve flows.
 7. The refrigerating apparatus according to claim 6, wherein the cylinder comprises: a cylinder body including at least one nozzle; and a cylinder flange that extends radially outward from the cylinder body.
 8. The refrigerating apparatus according to claim 7, wherein the frame comprises: a frame body that surrounds the cylinder body; a recessed portion which receives the cylinder flange; and a seating portion that faces a seating surface of the cylinder flange.
 9. The refrigerating apparatus according to claim 8, wherein the passage comprises a first passage between an outer circumferential surface of the cylinder flange and an inner circumferential surface of the recessed portion.
 10. The refrigerating apparatus according to claim 9, wherein the passage comprises a second passage between the seating surface of the cylinder flange and a seating surface of the frame.
 11. The refrigerating apparatus according to claim 10, wherein the passage comprises a third passage that extends into a space between an outer circumferential surface of the cylinder body and an inner circumferential surface of the frame body.
 12. The refrigerating apparatus according to claim 5, wherein the linear compressor comprises: a cylinder including a cylinder stepped portion on an inner circumferential surface thereof; a piston to reciprocate within the cylinder and including a piston stepped portion on an outer circumferential surface thereof, wherein a low pressure is formed between the piston stepped portion and the cylinder stepped portion when the piston moves in a first direction and a high pressure is formed between the piston stepped portion and the cylinder stepped portion when the piston moves in a second direction; an oil suction passage to allow oil to flow between the cylinder stepped portion and the piston stepped portion; and an oil discharge passage to allow the oil between the cylinder stepped portion and the piston stepped portion to be discharged outside of the cylinder.
 13. The refrigerating apparatus according to claim 1, wherein the non-azeotropic mixed refrigerant comprises a first hydrocarbon that is isobutane and a second hydrocarbon that is propane.
 14. The refrigerating apparatus according to claim 13, wherein the isobutane has a weight ratio of 76%≤isobutane≤87%.
 15. A refrigerating apparatus including a non-azeotropic mixed refrigerant, the refrigerating apparatus comprising: a linear compressor to compress the non-azeotropic mixed refrigerant; a condenser to condense the non-azeotropic mixed refrigerant compressed by the compressor; an expander to expand the non-azeotropic mixed refrigerant condensed by the condenser; and an evaporator to evaporate the non-azeotropic mixed refrigerant expanded by the expander, wherein a pressure difference (ΔP) of the non-azeotropic mixed refrigerant is in a range of 340 kPa<ΔP<624.7 kPa.
 16. The refrigerating apparatus according to claim 15, wherein the linear compressor comprises: a piston that reciprocates; and a cylinder to guide the piston, wherein the non-azeotropic mixed refrigerant of high pressure compressed by the piston is guided to an inner surface of the cylinder to cause an outer surface of the piston to be spaced from the inner surface of the cylinder.
 17. The refrigerating apparatus according to claim 15, wherein the non-azeotropic mixed refrigerant comprises at least two hydrocarbons, the at least two hydrocarbons comprising: at least one first hydrocarbon having an evaporation temperature of −12° C. or more at 1 bar; and at least one second hydrocarbon having an evaporation temperature of −50° C. or more and less than −12° C. at 1 bar, and wherein a gliding temperature difference is 4° C. or more.
 18. The refrigerating apparatus according to claim 17, wherein a weight ratio of the at least one first hydrocarbon is 50% or more.
 19. The refrigerating apparatus according to claim 15, wherein the linear compressor comprises: a piston that reciprocates to compress the non-azeotropic mixed refrigerant; and a cylinder to guide the piston, wherein oil pumped by a pressure difference of the non-azeotropic mixed refrigerant is present at contact surfaces between the piston and the cylinder.
 20. A refrigerating apparatus including a non-azeotropic mixed refrigerant, the refrigerating apparatus comprising: a compressor to compress the non-azeotropic mixed refrigerant; a condenser to condense the non-azeotropic mixed refrigerant compressed by the compressor; an expander to expand the non-azeotropic mixed refrigerant condensed by the condenser; and an evaporator to evaporate the non-azeotropic mixed refrigerant expanded by the expander and to provide cold air to an inner space of the refrigerating apparatus, wherein the compressor is selectively operable in an intermittent operation mode and a continuous operation mode, and wherein in the continuous operation mode, the compressor is controlled to operate continuously even when a temperature of the inner space of the refrigerating apparatus is within a target temperature range. 